The invention relates to a shaft seal for rotating shafts in turbo-machines, in particular but not exclusively to non-contacting shaft seals.
This type of shaft seal is often used with machinery for the pumping of gas (nitrogen, argon, hydrogen, natural gas, air, etc.) where the transmission of gas along the shaft needs to be prevented. Due to the high-pressure, high-speed machinery which is normally used, the shaft seals may be non-contact type seals, in order to reduce heat build up in the seals and the wear of the sealing parts.
Non-contacting operation avoids this undesirable face contact when the shaft is rotating above a certain minimum speed, which is often called a lift-off speed.
Non-contacting shaft seals provide advantages over seals where the sealing surfaces contact one another due to reduction in wear and the lower heat generation. Articles entitled xe2x80x9cFundamentals of Spiral Groove Non-contacting Face Sealsxe2x80x9d by Gabriel, Ralph P. (Journal of American Society of Lubrication Engineers Volume 35, 7, pages 367-375, and xe2x80x9cImproved Performance of Film-Riding Gas Seals Through Enhancement of Hydrodynamic Effectsxe2x80x9d by Sedy, Joseph (Transactions of the American Society of Lubrication Engineers, Volume 23, 1 pages 35-44) describe non-contacting seal technology and design criteria and are incorporated herein by reference.
As with ordinary mechanical seals, a non-contacting face seal consists of two sealing rings, each of which is provided with a very precisely finished sealing surface.
These surfaces are tapered-shape perpendicular to and concentric with the axis of rotation. Both rings are positioned adjacent to each other with the sealing surfaces in contact at conditions of zero pressure differential and zero speed of rotation. One of the rings is normally fixed to the rotatable shaft by means of a shaft sleeve, the other located within the seal housing structure and allowed to move axially. To enable axial movement of the sealing ring and yet prevent leakage of the sealed fluid, a sealing member is placed between the ring and the housing. This sealing member must permit some sliding motion while under pressure, therefore normally a top quality O-ring is selected for that duty. This O-ring is often called the secondary seal.
As mentioned above, to achieve non-contacting operation of the seal, one of the two sealing surfaces in contact is provided with shallow surface recesses, which act to generate pressure fields that force two sealing surfaces apart. When the magnitude of the forces resulting from these pressure fields is large enough to overcome the forces that urge seal faces closed, the sealing surfaces will separate and form a clearance, resulting in non-contacting operation. As explained in detail in the above-referenced articles, the character of the separation forces is such that their magnitude decreases with the increase of face separation. Opposing or closing forces, on the other hand, depend on sealed pressure level and as such are independent of face separation. They result from the sealed pressure and the spring force acting on the back surface of the axially movable sealing ring. Since the separation or opening force depends on the separation distance between sealing surfaces, during the operation of the seal or on imposition of sufficient pressure differential equilibrium separation between both surfaces will establish itself. This occurs when closing and opening forces are in equilibrium and equal to each other. Equilibrium separation constantly changes within the range of gaps. The goal is to have the low limit of this range above zero. Another goal is to make this range as narrow as possible, because on its high end the separation between the faces will lead to increased seal leakage. Since non-contacting seals operate by definition with a clearance between sealing surfaces, their leakage will be higher then that of a contacting seal of similar geometry. Yet, the absence of contact will mean zero wear on the sealing surfaces and therefore a relatively low amount of heat generated between them. It is this low generated heat and lack of wear that enables the application of non-contacting seals to high-speed turbomachinery, where the sealed fluid is gas. Turbocompressors are used to compress this fluid and since gas has a relatively low mass, they normally operate at very high speeds and with a number of compression stages in series.
As explained in the above-referenced articles, the effectiveness of the seal is largely dependent upon the so-called balance diameter of the seal. This is also true for contact seals. The pressure developed in the clearance, or gap, is balanced against the pressure existing against the back side of the fixed seal element. The closing forces, including a spring force and regions exposed to high and low (atmospheric) pressure, act to counter the opening force that creates the gap. An O-ring, commonly termed the secondary seal, divides the area at the back of the seal element into the regions of high pressure and low pressure. The O-ring slides backwards and forwards relative to the seal element, thus changing the relative size of the regions of pressure.
When pressure is applied from the outside diameter of the seal, reduction of the balance diameter results in a greater force pushing the two sealing faces together and so a smaller gap between the faces. Thus, less gas is leaked from the system.
During a typical period of operation, a turbocompressor is started and the power unit starts the shaft rotating. At the initial warm-up stage of operation, shaft speeds may be quite low. Typically, oil is used to support the shaft at its two radial bearings and one thrust bearing. Oil warms up in oil pumps and also accepts shear heat from compressor bearings. The oil together with process fluid turbulence and compression in turn warm-up the compressor. Once the full operating speed is reached, the compressor reaches in time some elevated equilibrium temperature. On shutdown, shaft rotation stops and the compressor begins to cool down. In this situation, various components of the compressor cool down at different rates and, importantly, the shaft contracts with decreasing temperature at a different rate than the compressor casing. These prior art secondary seal arrangements can be found for example in U.S. Pat. Nos. 4,768,790; 5,058,905 or 5,071,141. The term used often in the industry for this phenomenon is xe2x80x9cseal face hang-upxe2x80x9d. Often there is a very high leakage of process fluid the next time the compressor is restarted and often in such cases the seal will resist all attempts to reseal it. The seal must then be removed and replaced at a considerable cost in time and lost production.
U.S. Pat. No. 5,370,403 and EP-A-0,519,586 describe methods of reducing seal face hang up by attempting to prevent movement of the secondary seal.
The present invention seeks to provide a shaft seal with improved sealing characteristics in dynamically transient operation of the seal, as well as during pressurisation/depressurisation and stand-still conditions.
According to the present invention there is provided a shaft seal comprising a sealing element mounted co-axially to a rotary sealing ring to form a primary seal between opposed faces thereof to substantially prevent fluid flow across the primary seal from a high-pressure radial side to a low-pressure radial side, the sealing element being urged axially towards the rotary ring by a biasing means acting between a pusher sleeve connected to the sealing element and a seal housing with a secondary seal formed therebetween by a sealing member located in a channel in the part on the low-pressure radial side.
By providing the secondary sealing member in the part radially distant from the high pressure source, the primary seal quickly provides an effective barrier. Further, during start-up of the seal, the problem of seal face hang-up is greatly reduced or even eliminated. The inventor considers that the improvements due when the secondary sealing member slides, axially during start-up of the machine, the frictional engagement of the seal member with the pusher sleeve and housing surfaces, alters the balance diameter of the shaft seal to where the frictional engagement occurs. Consequently, due to the sealing member being located in the low pressure side, the balance diameter is altered towards providing increasing the closing force in the primary seal. Normally, the friction between the housing and pusher sleeve would be aligned along the equilibrium balance diameter of the shaft seal.
This invention is particularly applicable to non-contact seals.
Preferably, the pusher sleeve is a L-shaped part separate from the sealing element. The spring (biasing means) acts between the housing and one leg of the L-shaped sleeve parallel to the back face of the sealing element. The secondary sealing is formed between a face of the housing and the other leg of the L-shaped sleeve. Preferably the other leg of the sleeve is radially inward of the housing.
Advantageously, a further seal is provided by an O-ring between the pusher sleeve and the sealing element, wherein the further seal O-ring is located in a half dovetail channel. Normally, the further O-ring is provided between the pusher sleeve and the sealing element in a square cut or dovetail channel provided in the sleeve. The dove-tailed channel is often used so that the O-ring can be reliably installed in the seal. However, during shutdown of the seal, the seal is often blown out of the channel due to pressure build up in the dove-tailed channel which cannot otherwise be vented. The preferred half dove-tailed shape allows a limited movement of the O-ring allowing pressure to escape from the channel. Due to the very high cost associated with the seal strip down, this design is particularly advantageous.